Squeeze film bearing dampers are employed by high speed rotating machinery and particularly gas turbine engines to reduce and control vibration.
For some applications, squeeze film bearing dampers are accompanied by a flexible element in a parallel combination as a vibration isolator, so that the natural frequencies of the rotating system are artificially reduced to the extent that the speed range of high vibration amplitudes and transmitted forces are traversed well before the normal operating speed of the system is reached. In this manner, a squeeze film bearing damper acts simply as a device to reduce such amplitudes and transmitted forces to acceptable limits. Alternatively, squeeze film bearing dampers have been used alone between a bearing and its housing. Rotation of the outer race of any rolling element bearing is prevented by anti-rotation pins, called "dogs." In such an application, the role of the squeeze film bearing damper is solely one of damping with no intended effects on the natural frequencies of the rotating system. The resulting simpler mechanical design avoids the problem of fatigue in any introduced flexible element, and also reduces manufacturing costs.
In the rotor system of a gas turbine engine, a single ball bearing is typically used to carry thrust, while the remainder of the rotor is supported by two or more roller bearing assemblies.
The ball and roller bearing offer considerable load carrying capability but practically no damping. As is well known in the art, turbomachinery are prone to vibrations and dynamic loads caused by rotor and shaft unbalance and by self-excited whirl (i.e., dynamic instability). As a result, squeeze film bearing dampers have been widely used to damp shaft vibrations of gas turbine engines.
A conventional squeeze film damper (SFD) assembly 10 is shown in cross-section in FIG. 1. hose skilled in the art will appreciate that the assembly 10 is not shown to scale and is schematically represented for the sake of clarity. The assembly 10 includes a housing 12 having a cylindrical interior wall that defines a chamber 14, such that the chamber 14 has a uniform circular cross-section. The chamber 14 contains a bearing assembly 16, and the resulting annulus 18 defined between the housing 12 and bearing assembly 16 is filled with a semi-pressurized incompressible fluid, such as an oil. The bearing assembly 16 is typically a roller bearing whose inner race supports a shaft (not shown), the weight of which is indicated by the arrow "W" in FIG. 1. The outer race of the bearing assembly 16 has a cylindrical shape so that the housing 12 and bearing assembly 14 have a constant diametrical clearance, though the radial clearance and oil film thickness between the housing 12 and bearing assembly 16 will vary along the circumference of the bearing assembly 16 if the assembly 16 is not concentrically located within the chamber 14. In a conventional SFD, the outer race of the bearing assembly 16 is prevented from rotating relative to the housing 12, though radial and orbital movement of the bearing assembly 16 is permitted by providing much larger clearance between an anti-rotation device (e.g., pins or "dogs") and the outer race of the roller bearing.
During engine operation, the rotor is subjected to vibration which is transmitted to the bearing assembly 16 as an unbalance force F.sub.u, which is reacted by the oil film forces whose radial and tangential components are F.sub.r and F.sub.t respectively. Accordingly, vibration of the engine rotor produces vibratory and orbital movement of the bearing assembly 16 within the chamber 14, causing the bearing assembly 16 to exert a squeezing action on the fluid, moving and distributing the fluid throughout the annulus. This motion generates a hydrodynamic pressure field in the oil film. When the pressure is integrated along and around the SFD oil film forces, the reaction forces F.sub.r and F.sub.t are developed radially and tangentially, respectively. The tangential force F.sub.t provides the desired damping, while the radial force F.sub.r provides the stiffness which lifts the rotor within the SFD oil chamber. The latter is also called the lift force. These oil film forces are very non-linear in nature and increase with decreasing oil film thickness.
From the above, it can be appreciated that the ability of the SFD assembly 10 to damp vibration in a rotor system decreases as the bearing assembly 16 (and therefore the rotor) moves away from the wall of the chamber 14 and toward a concentric position with respect to the chamber 14. This phenomenon is the result of a positive oil pressure being created in a converging region 20 of the chamber 14, identified with a heavy outline in FIG. 1. The location of this region 20 relative to the line of centers is determined by the shaft rotation direction (.omega..sub.t, as a result of the orbital motion of the bearing assembly 16 within the chamber 14 induced by shaft vibration. The hydrodynamic pressure within the region 20 decreases as the radial gap diverges in the direction of shaft orbit (counterclockwise as depicted in FIG. 1). The remaining portion of the chamber 14 opposite the region 20 is generally at a negative pressure. As the radial gap between the bearing assembly 16 and the wall of the housing 12 increases, there is a decrease in the damping achieved by the squeeze effect of the rotor on the fluid film between the bearing assembly 16 and housing 12. Accordingly, the damping capability of the SFD assembly 10 is at a maximum when the bearing assembly 16 is immediately adjacent the chamber wall because of the greater positive fluid pressure generated within the converging annular gap between the bearing assembly 16 and wall of the chamber 14.
Rotating shafts supported in the manner shown in FIG. 1 are often subject to what is termed the "jump phenomenon" between two stable vibration modes. The jump phenomenon is illustrated in FIG. 2 with a graph plotting vibration amplitude versus frequency, with the frequency .omega..sub.1 shown as having a bistable response between amplitudes A.sub.1 and A.sub.2. These amplitudes are. both possible and stable, and the rotor can "jump" between these amplitudes in response to small external disturbances. However, in rotating machinery such as the rotor of a gas turbine engine, only one stable amplitude is desirable to promote the service life of the rotor and other engine components. Some oil supply pressure control aschemes involving check valves may reduce the possibility of bistable rotor operation, but these designs are complicated and expensive, and are also susceptible to unexpected malfunctions.
Thus, it would be desirable if a turbo machine shaft support system were available by which improved shaft vibration damping was achieved without contributing to bistable vibration and without complicated hardware and control requirements.